The invention relates to a drag damper designed to be fitted between the hub of a rotary-wing aircraft rotor and a flapping mass, which comprises one of the blades of the rotor and a device connecting this blade to the hub so as to damp the angular drag movements of said flapping mass with respect to the hub, i.e. the angular deflections of the blade, and more globally of the corresponding flapping mass, about their drag axis which is substantially parallel to the axis of rotation of the rotor;
The rotor is more particularly a helicopter main rotor subject to the instability phenomena known as “ground resonance” and “air resonance”, although a conventional tail rotor may also be equipped with drag dampers according to the invention.
On rotors of the hinged type, the device connecting a blade to the hub may be arranged as a means of securing the blade and hinging it to the hub, when the blade is connected by its root, possibly in the form of a fork, to the hub, or as a device which is substantially radial (relative to the rotor axis) generally termed a cuff, and fitted with yokes at the ends to be connected to the blade root on the one hand and on the other to means of securing and hinging, such as a spherical laminated stop, itself connecting it to the hub, while on rotors of the semi-rigid type, this connecting device may be a flexible torsion arm, at the blade root, and surrounded by a torsionally rigid cuff integral with the blade root for controlling the blade in pitch, which is connected and hinged to the hub by this flexible torsion arm.
Numerous different embodiments of drag dampers are known, particularly dampers which are hydraulic, hydro-pneumatic, laminated with at least one layer of visco-elastic material stressed between two rigid fittings, or comprising combinations of these different means, these drag dampers comprising means of elastic return of defined stiffness and damping, when they are fitted to helicopter main rotors, to combat the resonance phenomena mentioned above.
It is a well-known practice to design into helicopter rotor blades and therefore into the corresponding flapping masses a natural drag frequency, also termed first drag mode or natural drag mode, which is different from the nominal rotation frequency at which the rotor is designed to be driven.
More generally, to avoid in particular fatigue problems resulting from the dynamic stresses in the blades and the fuselage, and problems of vibration levels in the fuselage, it is essential to position correctly the natural frequencies of the blades in flapping, torsion and drag relative to the nominal rotation speed of the rotor and its harmonics (multiples).
This results from the fact that a helicopter rotor constitutes a powerful vibration generator. Because of the variable angles of incidence and speeds of rotor blades and also of helicopters, alternating loads of aerodynamic origin are developed notably in the blades of rotors, and give rise in the latter to stresses as well as reactions on the attachments, particularly of the blades to the hubs. From this there result alternating loads and moments on the rotor heads, and the development of high vibration and stress levels in fuselages. The response of each blade, the stresses to which this blade is subjected and the loads which this blade transmits to the hub at the centre of the rotor are all the greater as at least one natural frequency of the blade (in drag, flapping and torsion) is close to the rotation frequency Ω of the rotor or one of its harmonics nΩ (where n is a whole number).
The dynamic characteristics of the rotor blades are therefore chosen to obtain suitable dynamic matching by ensuring that their natural vibration frequencies in flapping, drag and torsion are correctly positioned relative to the nominal rotation frequency Ω of the rotor and its multiples nΩ, which is why it is necessary to observe certain simple rules for positioning the frequencies, and in particular two essential rules, the first of which is to avoid positioning a natural vibration frequency in flapping, drag or torsion on or very close to a harmonic of the rotation speed nΩ (where n≧1), and the second to endeavour as far as possible to position only one of these three natural frequencies between two adjacent harmonics nΩ and (n+1) Ω of the rotation speed in order to avoid coupling.
In addition to these two essential rules, it is imperative to follow recommendations proper to each type of deformation in flapping, drag or torsion.
Concerning the recommendations relating particularly to the drag modes for hinged or semi-rigid (semi-hinged) rotors, the first drag mode (or natural drag frequency) is at the origin of “ground resonance” and “air resonance” problems due to coupling with modes of the helicopter structure.
On a rotor with blades hinged in drag, the angular frequency or pulsatance of the first drag mode is given by the expression:                ω      ⁢                           ⁢      δ        =                  Ω        ⁡                  (                                                    e                ·                m                            ⁢                                                           ⁢              δ                                      1              ⁢              δ                                )                            1        /        2            where e is the drag eccentricity of each blade, mδ is the static moment of the flapping mass (blade+device connecting it to the hub) relative to the hinge (drag axis) and Iδ is the inertia of the flapping mass relative to this drag hinge.
On a semi-rigid rotor, the first drag mode of a blade or flapping mass depends on the characteristics not only of the blade or flapping mass but also of the hub. The pulsatance of the first drag mode is then given by the expression:                ω      ⁢                           ⁢      δ        =                  Ω        ⁡                  (                                                                      e                  ·                  m                                ⁢                                                                   ⁢                δ                                            1                ⁢                δ                                      +                                          k                ⁢                                                                   ⁢                δ                                                              Ω                  2                                ⁢                1                ⁢                                                                   ⁢                δ                                              )                            1        /        2            where kδ is the stiffness of the drag damper fitted between the blade or corresponding flapping mass and the hub of the rotor.
The positioning of the first drag mode of a blade or of the corresponding flapping mass depends upon the modes of the helicopter structure (fuselage mass, inertia, stiffness of the landing gear and of any tyres which may be fitted to it), these modes of the structure being generally determined by specific tests, adjustment of the first drag mode being obtained by altering the term kδ representing the stiffness of the drag damper.
As a general rule, as the upper limit of the first drag mode ωδ, a value close to three-quarters of the nominal rotation frequency Ω of the rotor is taken, so as not to introduce excessively high stresses in the blades of the rotors. One of the other two natural vibration frequencies (in flapping and in torsion) of the blade or flapping mass is placed between Ω and 2Ω, and the other, as far as possible, between 2Ω and 3Ω.
For this reason, when the rotor is started up or stopped, and also at the end of a landing by the helicopter in autorotation, the instantaneous speed of rotation of the rotor intersects the resonance frequency in drag situated below the nominal speed. Because of this, and also because of the fairly large range of variations in rotor rotation speeds which are authorised for helicopters in flight, it is necessary to increase damping at the natural vibration frequencies of the blades in drag, and possibly to reduce this natural frequency by means of drag dampers, which is the reason why these dampers are also termed frequency adapters, the aim being that the blades should be sufficiently damped in drag to avoid going into resonance.
The invention relates more specifically to a drag damper of the general type comprising a tubular damper body in which a piston moving integrally with a damper rod and slidable axially is fitted, and in the damper body delimits two opposing variable volume working chambers, filled with a fluid of which volumes are transferred, by restriction of the fluid by at least one flow-restriction port arranged between the piston and the body and/or in the piston, between the working chambers when the piston moves in the damper body, upon which elastic means bear and load the piston and/or the rod so as to return the rod-piston assembly towards a neutral position in the body.
A drag damper of this kind, known in particular from FR 2 063 969, may be hinged on the one hand to a fixed point on the hub or on a bracket connected to the rotor hub, by means connecting the body or one end, external to said body, of the damper rod and on the other to a fixed point on the corresponding flapping mass, at the blade root of this flapping mass or on a device connecting this blade to the hub, by means connecting the end of the rod external to the body or the damper body respectively.
As the hinging point of one end of the drag damper on the hub or a bracket fixed to the hub is situated between the blade on which the drag damper is hinged at its other end and an adjacent blade, the stiffness of the damper introduces an equivalent angular stiffness, opposed to the angular deflections of the blade relative to the hub about its drag axis. It is thus possible to increase the natural frequency of the blades in drag to escape from the two resonance phenomena mentioned above, with additional damping at the natural drag frequency ωδ of the blade when the phenomena of air and ground resonance occur.
However, it is known these phenomena rarely appear during the life of a helicopter. Most of the time, the drag dampers are subject to forced excitation at the rotation frequency Ω of the rotor, on which the drag dampers dissipate energy to no purpose.
The mean power dissipated in a drag damper of a rotor can be expressed by the following relation: Pd=π·K″·f·Xe2, where K″ is the dissipative stiffness of the damper, f the frequency of the movement applied to the damper (axial movement of the rod-piston assembly in the cylinder) and Xe is the movement of said rod-piston assembly associated with frequency f.
For example, for drag dampers of the type presented above fitted to a four-bladed main rotor of a helicopter with a weight of about 8 to 10 tonnes, a comparison of the energy dissipated on the forced excitation at the rotation frequency Ω with that which is dissipated on the natural frequency in drag ωδ of the blades gives the following results.
For the same dissipative stiffness K″ of 400 daN/mm, the forced excitation at Ω with a frequency f of 4.5 Hz corresponds to an associated displacement Xe of 4 mm, that is to say a dissipated power of 900 W, whereas for the natural drag mode ωδ at a frequency f of 2 Hz, causing an associated displacement Xe of 1 mm, there corresponds a dissipated power of 30 W.
Each drag damper therefore dissipates 97% of its energy on forced excitation at Ω. Under these conditions of use, this energy is dissipated to no purpose, which entails substantial component fatigue, not only of the drag damper but of the means connecting it to the hub and to the flapping mass, and wasted weight due to oversizing of these parts.
On the helicopter with a weight of about 8 to 10 t considered, the forces applied to the drag dampers upon forced excitation at Ω are very high, which may cause incidents in service such as cracks in the yokes connecting the drag dampers to the flapping masses, damage to the drag damper spindles at the end where there are connected to the hub and also damage to the fittings connecting the drag dampers to the hub, and rapid deterioration of the ball joints used in these connection devices.
Moreover, FR 2 737 271 makes known a damper which senses acceleration and can be used in numerous applications in which it is necessary to damp elastic systems, and particularly as a damper for suspension systems, shock and yaw control, in transport and industry.
A damper of this kind comprises a tubular body in which a piston moving integrally with a damper rod and slidable axially is fitted, and in the body delimits two opposing variable volume working chambers connected to each other by at least one bypass channel, the length of which is very much greater than a main dimension of its cross-section, which is itself very much smaller than the cross-section of the body, a fluid filling at least the two working chambers in the body and the bypass channel, which can be made in the wall of the body, or in the piston, having the form of a spiral, helix or one or more concentric arcs of circle. When the piston is moved in the body, fluid is expelled and moves from one working chamber to the other using the bypass channel. By calculation, taking account of the law of conservation of fluid flows, the force to which the piston is subjected and accelerations of the piston on the one hand and of the column of fluid in the bypass channel on the other, it can be shown that the force to which the piston is subjected is proportional to the virtual fluid mass, similar to inertia, which can be maximised by adjusting the mass per unit volume of the fluid, the length of the bypass channel and above all the ratio between the cross-section of the body and the cross-section of the bypass channel. As viscous damping increases with the reduction in the cross-section and increase in the length of the bypass channel, choice of a low-viscosity fluid is indicated if the function of locking the rod-piston assembly in the cylinder by fluid inertia is to be favoured above the viscous damping function.
Moreover, EP 0 183 039 and GB 2 111 171 make known the practice of adjusting the damping force of a hydraulic damper of conventional type by filling its working chambers with an electro-rheological fluid circulating in a bypass channel connecting the two working chambers and subjected, in this bypass channel, to a variable electrostatic field enabling the viscosity of the fluid to be varied according to the signals from detectors such as velocity, acceleration, load detectors etc.